Centrifugal pump with hydrodynamic bearing and double involute

ABSTRACT

A centrifugal pump with a hydrodynamic bearing and a double involute is disclosed. Some embodiments may include an impeller housing comprising a fluid entrance to allow fluid to enter and an impeller located within the impeller housing, the impeller comprising a plurality of impeller blades, a plurality of fluid channels between the impeller blades, and a motor magnet. The impeller may rotate within the impeller housing about a pump centerline in response to an electromagnetic field and the fluid channels may each allow fluid to pass through when the impeller rotates. Embodiments may also include one or more hydrodynamic bearings positioned between the impeller and the impeller housing to support generated loads and a double involute coupled with the impeller housing and positioned to receive fluid exiting the plurality of fluid channels. Further embodiments may include a motor stator to generate the electromagnetic field. Other embodiments are disclosed and claimed.

FIELD

The present invention is in the field of centrifugal pumps. More particularly, the present invention relates to a centrifugal pump with a hydrodynamic bearing and a double involute.

BACKGROUND

Electronic devices generate heat during operation. Device designers often utilize thermal management to keep temperature-sensitive elements of an electronic device within a prescribed operating temperature. Failure to properly cool an electronic device can result in overheating, which may cause a reduction in device service life, device failure, or a reduction in operating performance. Historically, designers have cooled electronic devices using natural convection by strategically locating openings in the device packaging or case to allowed warm air to escape and cooler air to be drawn in. The advent of high performance electronic devices such as processors, however, now requires more sophisticated thermal management. Increasing electronic device performance often results in an increase in the heat generated by the device and often results in a smaller size for the electronic device, both conditions of which increase the amount of thermal energy that needs to be handled. As electronic device designs continue to increase in sophistication, these problems will be exacerbated and the need for improved thermal management will continue to increase.

One thermal management solution for high performance processors or other computer system components is the use of liquid cooling. One method of liquid cooling of components is to use a cold plate thermally coupled to the component. In this solution, a pump may pump cooling fluid through the cold plate, allowing heat to be transferred from the component to the cooling fluid through the cold plate, after which the heat is removed from the cooling fluid via a heat exchanger and then returned to the cold plate. Liquid cooling using a cold plate can be more effective than solid conduction cooling methods and can also provide additional flexibility in the size and location of the heat exchanger, as the system can pump the heated fluid to a heat exchanger located in a more desirable location. While a liquid cooling system can be effective at cooling high performance components, it can be more expensive and complicated than previous methods. Because of the cost and complexity of liquid cooling systems, liquid cooling is typically only used on higher end systems. The cost and complexity of pumps to move cooling fluid through a liquid cooling system is a significant part of the cost and complexity of the entire liquid cooling system. Reducing the cost and complexity of liquid cooling pumps can therefore make liquid cooling solutions for heat-generating components suitable for more systems.

BRIEF DESCRIPTION OF THE DRAWINGS

Advantages of the invention will become apparent upon reading the following detailed description and upon reference to the accompanying drawings in which like references may indicate similar elements:

FIG. 1 depicts a side cut-away view of a centrifugal pump with a double involute and hydrodynamic bearings according to one embodiment;

FIG. 2 depicts a side cut-away view of a hydrodynamic thrust bearing of the centrifugal pump of FIG. 1 according to one embodiment;

FIG. 3 depicts a cut-away plan view of a double involute of the centrifugal pump of FIG. 1 according to one embodiment; and

FIG. 4 depicts a flowchart of an embodiment to pump cooling fluid in a cooling system.

DETAILED DESCRIPTION OF EMBODIMENTS

The following is a detailed description of example embodiments of the invention depicted in the accompanying drawings. The example embodiments are in such detail as to clearly communicate the invention. However, the amount of detail offered is not intended to limit the anticipated variations of embodiments; on the contrary, the intention is to cover all modifications, equivalents, and alternatives falling within the spirit and scope of the present invention as defined by the appended claims. The detailed descriptions below are designed to make such embodiments obvious to a person of ordinary skill in the art.

Generally speaking, a centrifugal pump with a hydrodynamic bearing and a double involute is disclosed. Some embodiments may include an impeller housing comprising a fluid entrance to allow fluid to enter and an impeller located within the impeller housing, the impeller comprising a plurality of impeller blades, a plurality of fluid channels between the impeller blades, and a motor magnet. The impeller may rotate within the impeller housing about a pump centerline in response to an electromagnetic field and the fluid channels may each allow fluid to pass through when the impeller rotates. Embodiments may also include one or more hydrodynamic bearings positioned between the impeller and the impeller housing to support generated loads and a double involute coupled with the impeller housing and positioned to receive fluid exiting the plurality of fluid channels. Further embodiments may include a motor stator to generate the electromagnetic field.

Another embodiment comprises a method for pumping cooling fluid. Some embodiments of the method may include receiving a cooling fluid into an impeller and driving the impeller to rotate about an axis to force the cooling fluid to exit the impeller. Embodiments may also include receiving the exiting cooling fluid at two or more involutes and increasing a static pressure of the cooling fluid in the two or more involutes. Embodiments may also include reacting any loads generated by the two or more involutes and rotating impeller with one or more hydrodynamic bearings. Further embodiments may include after increasing the static pressure of the cooling fluid, passing the cooling fluid to a cooling system.

The disclosed system and methodology may advantageously provide for a centrifugal pump with two or more involutes, such as a double involute, and hydrodynamic bearings. The use of a two or more involutes may reduce radial and moment loads on the impeller and centrifugal pump and may thus allow for the use of hydrodynamic bearings instead of solid element bearing technologies, potentially reducing the cost and complexity of the centrifugal pump. The combination of a multiple involutes with hydrodynamic bearings may also allow for the use of plastic injection molded parts for the centrifugal pump, potentially further reducing the cost of the pump. Reduced cost and complexity for centrifugal pumps may allow for use of centrifugal pumps and liquid cooling in a more diverse set of circumstances and for more types of systems.

Various embodiments of the present invention provide systems and methods for pumping fluid. The following description provides specific details of certain embodiments of the invention illustrated in the drawings to provide a thorough understanding of those embodiments. It should be recognized, however, that the present invention can be reflected in additional embodiments and may be practiced without some of the details in the following description. In other instances, well-known structures and functions have not been shown or described in detail to avoid unnecessarily obscuring the description of the embodiments of the invention. While specific embodiments will be described below with reference to particular configurations and systems, those of skill in the art will realize that embodiments of the present invention may advantageously be implemented with other substantially equivalent configurations and/or systems.

Turning now to the drawings, FIG. 1 depicts a side cut-away view of a centrifugal pump with a double involute and hydrodynamic bearings according to one embodiment. The centrifugal pump 100 of FIG. 1 includes an impeller 106 having a motor magnet 108 housed within an impeller housing 102. A motor stator 104 may be located in or outside the impeller housing 102. When electric power is applied to the motor stator 104, the motor stator 104 creates a magnetic field that may drive the motor magnet 108 of the impeller 106, causing the impeller 106 to rotate within the impeller housing 102 and about a pump centerline 110. Cooling fluid may enter the impeller housing 102 through, for example, a fluid entrance 126. As the axial flow input enters the impeller housing 102 through the fluid entrance (and along the pump centerline 110), it flows into the rotating impeller 106. The rotational speed of the impeller 106 creates a centrifugal force that propels the fluid through fluid channels 114 formed by and between a plurality of impeller blades 112. The fluid exiting the fluid channels 114 may be collected by a double involute 132 at two involute tongues 130. The double involute 132 (as described in more detail in relation to FIG. 3) may then convert the dynamic pressure of the fluid to a static pressure, generating the pressure difference used by the centrifugal pump 100 to drive the cooling fluid. The double involute 132 may generate equal and opposite loads at each involute which cancel out to create a relatively small net radial load. With a double involute 132, the generated loads are equal and opposite when the impeller 106 has an even number of impeller blades 112 evenly distributed around the impeller 106 and hence an impeller blade 112 passes both involute tongues 130 at the same time. A double involute 132 may thus provide a more balanced and lower radial load on the bearings of the centrifugal pump 100 when compared to a single involute.

The centrifugal pump 100 of the disclosed embodiments may utilize hydrodynamic bearings between the impeller 106 and impeller housing 102 to support the rotating impeller 106 and its associated loads within the impeller housing 102. Hydrodynamic bearings (which are described in more detail in relation to FIG. 2) may rely on the gap between a stationary surface (i.e., the inside surface of the impeller housing 102) and a moving surface (i.e., the outside surface of the rotating impeller 106) and viscous effects of its constituent bearing film to handle the net radial force and moment associated with the development of mass flow and static pressure on the impeller blades 112. The hydrodynamic bearings may include one or more journal bearings 122 positioned between the impeller 106 and the impeller housing 102, such as near the motor stator 104, and one or more thrust bearings 124 positioned between the impeller blades 112 and the impeller housing 102. Hydrodynamic bearings may be less expensive than other types of bearings (such as roller bearings, ball bearings, needle bearings or other solid element bearings) but often have less load-bearing capacity than the more expensive bearings. The reduced loads resulting from the double involute 132 may advantageously allow the use of hydrodynamic bearings in the centrifugal pump 100 instead of more expensive and higher capacity bearings, providing for a less expensive centrifugal pump 100.

The impeller housing 102 of the depicted embodiment is symmetrical about the pump centerline 110, completely enclosing the impeller 106 while allowing the passage of fluid into the impeller housing 102 through the fluid entrance 126. The fluid entrance 126 may be an opening (or multiple openings) in the impeller housing 102 or other component (such as a tube or pipe) that allows passage of cooling fluid into the impeller housing 102. The impeller housing 102 may be made of any type of material, including metal or plastic, and may be of any shape adapted to partially or fully enclose the impeller 106. In one embodiment, the impeller housing 102 may be manufactured using plastic injected part manufacturing methods. The impeller housing 102 may be oriented vertically (as depicted in FIG. 1) with a gravitational force oriented downward, horizontally with the pump centerline 110 perpendicular to the gravitational force, or in any other direction. The impeller housing 102 may be part of a larger system or may be coupled to other components, such as a cold plate for a liquid cooling system.

The impeller 106 may include an impeller body 116 in addition to the motor magnet 108. Similarly to the impeller housing 102, the impeller body 116 may be made of any type of material, including metal or plastic, and may be injection molded plastic in one embodiment. The motor magnet 108 may be attached to the impeller body 116 in any fashion such that a rotational force applied to the motor magnet 108 also rotates the impeller body 116 (and thus the impeller 106 as a whole). This allows the impeller 106 to be magnetically coupled with the motor stator 104 through the wall of the impeller housing 102. This eliminates the need to have a shaft or other physical coupling mechanism to connect to drive the impeller 106. This may reduce cost and complexity, as a shaft through the impeller housing 102 to drive the impeller 106 would require, for example, rotating seals, which may be expensive and prone to leakage. In one embodiment, the motor stator 104 is outside the impeller housing 102 and shaped as a concentric circle around the outside of the impeller housing 102. For example, the motor stator 104 may be one or more laminated steel sheets with copper wires wound on it and wrapped around the impeller housing 106 that generates a magnetic field when a direct current (DC) charge is applied to the copper wire. Other motor stator 104 designs may also be used, including solid motor stators 104 or a motor stator 104 that is partially or fully integrated within the wall of the impeller housing 102.

The impeller body 116 and motor magnet 108 may be configured in any way. In one embodiment, for example, the motor magnet 108 may be positioned outside of the impeller body 116 so that the motor magnet 108 is closer to the motor stator 104. This embodiment may maximize the magnetic force created by the motor magnet 108 and motor stator 104 as the motor magnet 108 will be closer to the motor stator 104 and thus in a more powerful part of the motor stator's 104 magnetic field when it is powered. In another embodiment, the motor magnet 108 may be positioned on the inside of the impeller body 116 so that it is closer to the pump centerline 110. In other embodiments, the impeller body 116 may fully or partially enclose the motor magnet 108. In an alternative embodiment, the motor magnet 108 may serve as the impeller body 116, eliminating the need for a separate motor magnet 108 and impeller body 116.

As described previously, the rotational speed of the impeller 106 creates a centrifugal force that propels cooling fluid through the fluid channels 114 between the plurality of impeller blades 112 of the impeller 106. The impeller blades 112, which may be part of the impeller 106, may be any shape or size. Likewise, the fluid channels 114 formed between the impeller blades 112 to allow passage of cooling fluid through their length may be any size or shape suitable to allow passage of cooling fluid. The centrifugal force created when the impeller 106 pushes cooling fluid through the fluid channels 114 from the inside of the impeller blade 112 towards the impeller housing 102 to the outside of the impeller blade 112.

The cooling fluid exiting the fluid channels 114 of the impeller blades 112 may be collected by an involute. An involute may be any geometry that collects fluid exiting the impeller blades 112 and efficiently increases the static pressure of the fluid before it exits the involute. The involute may accomplish the increased static pressure by converting the dynamic pressure resulting from the circumferential velocity of the fluid and converting it to static pressure of the cooling fluid. The involute may advantageously be a double involute 132 having two involute channels 134 over at least part of its length. The double involute 132 may be positioned on the exterior of the impeller housing 102 and wrapped around the circumference of the impeller housing 102. The double involute 132 may have two involute tongues 130 (which may also be known as cutwaters). The involute tongue 130 may represent the closest point of the double involute 132 to the fluid channel exit 114 and as such is the point where the double involute 132 interacts with the cooling fluid exiting the impeller 106.

The conversion of the velocity of the cooling fluid to an increase in static pressure of the fluid in the involute 132 may result in a reactionary force on the impeller 106. Due to the close proximity of the involute tongue 130 to the impeller blades 112, the involute tongue 130 may generate a reaction force on the impeller blades 112, thus creating a radial and tangential reaction force in the impeller 106 itself. The tangential reaction force may be overcome with the torque capability of the motor (the combined motor stator 104 and motor magnet 108). The net radial load (both its magnitude and direction) may result from the integration of the radial reaction forces generated by the double involute 132 over its entire circumference of 360 degrees. By employing two involute channels 134 and involute tongues 130 that are approximately 180 degrees apart, the double involute 132 may substantially balance the impeller loading, resulting in smaller net radial loads when compared to single involute designs. To accomplish this, the double involute 132 may generate equal and opposite loads at each involute which cancel out to create a relatively small net radial load. To help maintain a low net radial load, the impeller 106 may have an even number of impeller blades 112 evenly distributed around the impeller 106, resulting in an impeller blade 112 passing both involute tongues 130 simultaneously and resulting in generated loads that are equal and opposite. Single involute designs may result in such high radial loads that hydrodynamic bearings are not viable and more expensive mechanical bearing approaches must be employed. A double involute 132 may thus provide a more balanced and lower net radial load on the bearings of the centrifugal pump 100 when compared to a single involute.

In an alternative embodiment, the centrifugal pump 100 may have more than two involutes instead of a double involute 132. In this embodiment, the number of impeller blades 112 may be matched to the number of involutes to balance the net radial loads. A three involute centrifugal pump 100, for example, may have a plurality of impeller blades 112 in a multiple of three (e.g., 3, 6, 9, 30, etc.) in order to substantially balance the loads. Similarly, a five involute centrifugal pump 100 may have impeller blades 112 in multiples of five (e.g., 5, 10, etc.) to properly balance the loads. A centrifugal pump 100 with an even number of involutes (e.g., 2, 4, 6, etc.) may have an even number of impeller blades 112 to provide load balancing. A centrifugal pump 100 with the number of involutes and the number of impeller blades 112 matched to achieve load balancing may have significantly less net radial loads than a single involute design, regardless of the number of impeller blades 112. While a double involute centrifugal pump 100 is depicted in the Figures and described herein, one skilled in the art will recognize that other multiple involute designs may also be used.

Hydrodynamic bearings such as the journal bearing 122 and thrust bearing 124 may support the loads generated by the operation of the centrifugal pump 100. The journal bearing 122 may support the rotating impeller 106 and motor magnet 108 and a combination of the journal bearing 122 and thrust bearing 124 may support the net radial load and force moment associated with the development of mass flow and static pressure on the impeller blades 112. The thrust bearing 124 may also help balance component gravity loads (depending on the orientation of the centrifugal pump 100) as well as support the moment loads generated by the offset between the journal bearing 122 surface and the impeller 106 height difference (where radial load is generated). The journal bearing 122 may be created or positioned in the gap between the motor magnet 108 and/or impeller body 116 and the impeller housing 102 in the region near the motor stator 104 (or the surface of the motor stator 104 itself). The journal bearing 122 may include a hydrodynamic fluid located in the gap between the rotating cylinder (the impeller 106) and the stationary cylinder (the motor stator 104 or impeller housing 102). The journal bearing 122 may be created or positioned either on the outside of the rotating cylinder, the inside of the rotating cylinder, or both. The thrust bearing 124, on the other hand, may be created or positioned in the gap between the bottom of the impeller 106 (e.g., the disk surface of the impeller 106) and the impeller housing 102 and/or the gap between the top of the impeller blades 112 and the impeller housing 102.

The journal bearings 122 and thrust bearings 124 are hydrodynamic in nature and their performance may typically be dominated by the viscous effects between the stationary (i.e., the impeller housing 102) and moving surfaces (i.e., the impeller 106). The effectiveness of the hydrodynamic lubrication may be closely coupled to surface geometry and interfaces (i.e., gaps between surfaces). The effectiveness and load capability of the journal bearing 122, for example, may be related to the eccentricity of the inner cylinder (i.e., the shape of the impeller 106), the viscosity of the fluid in the gap, the speed of the rotating impeller 106, the radius of the inner cylinder, and the radial clearance between the two cylinders (i.e., the gap), or other factors. For example, a reduction in the gap of the journal bearing 122 may improve the bearing performance. A gap reduction may also improve the efficiency of the motor by bringing the motor magnet 108 and motor stator 104 closer together and increasing the strength of the generated magnetic field on the motor magnet 108 and reducing the unused field, or magnetic leakage, between the two components. Improved motor efficiency may allow for less complicated control circuitry or less windings. Reduction of the gap, on the other hand, may also increase the torque losses in the motor due to the increase in shear losses in the fluid, and an optimal gap may exist based on balancing the different factors and the particular configuration. Thrust bearings 124 are impacted by similar factors as journal bearings 122, such as by the geometry of the bearing surfaces, the speed of rotation, etc.

FIG. 2 depicts a side cut-away view of a hydrodynamic thrust bearing of the centrifugal pump of FIG. 1 according to one embodiment. In FIG. 2, the thrust bearing 124 is positioned between a housing inner surface 206 of the impeller housing 102 and a thrust bearing surface 204 of the impeller body 116 part of the impeller blade 112. The distance between the impeller body 116 and the impeller housing 102 that forms the thrust bearing 124 is the thrust bearing gap 202. As described in relation to FIG. 1, a fluid channel 114 may be positioned on the opposite side of the impeller body 116 from the thrust bearing 124. While FIG. 2 is described as a thrust bearing 124 herein, the description of FIG. 2 is equally applicable to a journal bearing 122 with, for example, a motor magnet 108 replacing the impeller blade 112 or impeller body 116 in the description.

As described previously, the size of the thrust bearing gap 202 may impact the performance of the thrust bearing 124. A smaller thrust bearing gap 202 may result in improved load capacity of the thrust bearing 124 while larger thrust bearing gaps 202 may result in reduced capacity. In one embodiment, the geometry and tolerances associated with the bearing surfaces (thrust bearing surface 204 and housing inner surface 206) may be such that the thrust bearing gap 202 does not grow so large that the thrust bearing 124 cannot support the required loads. Tighter manufacturing tolerances of the housing inner surface 206 and thrust bearing surface 204, as examples, and reduced size of the thrust bearing gap 202 may therefore provide for more consistent and predictable load bearing capability as well as increased capability. Close manufacturing tolerances may also prevent gross instabilities in the position of the impeller 106 and therefore load fluctuations on the impeller 106, as the loading on the impeller blades 112 may be dependent on their distance in interaction with the double involute 132 geometry.

In one embodiment, a journal bearing 122 may be constructed with less expensive materials while still maintaining sufficiently close tolerances. The close tolerances may include both dimensional geometric aspects of the journal bearing 122 cylinders (i.e., diameters and eccentricity of the hole and cylinder). Close tolerances may be achieved using cost-effective plastic injected part manufacturing by providing support to the resulting thermoplastic walls by using components that act as a stabilizing backbone. For example, the journal bearing thermoplastic surface (equivalent to the housing inner surface 206 of the thrust bearing 124) may be stabilized by a metal motor stator 104. Similarly, the journal surface (equivalent to the thrust bearing surface 204 of the thrust bearing 124) may be stabilized with the motor magnet 108 ring. The composite construction may allow for more stable dimensional control and less geometric departure than simple plastic injection molded surfaces.

FIG. 3 depicts a cut-away plan view of a double involute of the centrifugal pump of FIG. 1 according to one embodiment. In the embodiment of FIG. 3, the centrifugal pump 100 includes a double involute 132 around an impeller 106 that is adapted to rotate clockwise (in FIG. 3) about a centerline 110. The impeller 106 may include a plurality of impeller blades 112 with fluid channels 114 to transport cooling fluid from the impeller 106 to the double involute 132 at a high velocity. The walls of the impeller blades 112 depicted in FIG. 3 are shown with no thickness for purposes of clarity. The double involute 132 may collect the high velocity fluid exiting the fluid channels 114 and may efficiently increase the static pressure of the fluid by the geometry of the double involute 132, resulting in a higher pressure fluid for use in a cooling system.

The double involute 132 of the disclosed embodiments includes two involute channels 134 each beginning at an involute tongue 130 and ending at an involute throat 302. Fluid may enter each involute channel 134 at the involute tongue 130 and travel the length of the involute channel 134 between its walls until exiting the double involute 132 at the involute throat 302. The involute tongues 130 are the start of the involute geometry and may represent the closest point that the involute wall interacts with the fluid exiting the impeller 106. The involute throat 302 may be of any length or configuration and may fluidly connect to a cooling system. Due to its close proximity, the involute tongue 130 generates the largest reaction force on the impeller blades 112, thus creating a radial and tangential reaction force in the impeller 106 itself. This tangential force may be overcome with the motor's torque capability. The integration of the radial reaction forces generated by the involute walls over their entire length generates the net radial load, including both its magnitude and direction.

The design of the double involute 132 helps reduce net radial loads in comparison to single involute designs. In the depicted double involute 132 design, the two involute tongues 130 (labeled ‘A’ and ‘B’ in FIG. 3) are positioned opposite each other and approximately 180 degrees apart. This may substantially balance the impeller 106 loading and results in a relatively small net radial load. The outside involute channel 134, which begins at involute tongue 130 ‘A’, may receive fluid exiting impeller blades 112 positioned between involute tongue 130 ‘A’ clockwise to involute tongue 130 ‘B’ (from approximately the 9 o'clock position to approximately the 3 o'clock position). The fluid gathered by the outer involute channel 134 then travels around the channel until it exits at the involute throat 302, giving it approximately 180 degrees of travel within the outer involute channel 134. The inner involute channel 134, which begins at involute tongue 130 ‘B’, may receive fluid exiting impeller blades 112 positioned between involute tongue 130 ‘B’ clockwise to involute tongue 130 ‘A’ (from approximately the 3 o'clock position to approximately the 9 o'clock position). The fluid gathered by the inner involute channel 134 travels around the channel until it exits at the involute throat 302, giving it up to 180 degrees of travel within the inner involute channel 134. By employing two involute channels 134 and associated geometry that are approximately 180 degrees apart the impeller 106 loading may advantageously be directly balanced and the net radial load reduced when compared to single involute designs. Positioning of the involute tongues 130 in separate halves of the double involute 132 may be considered positioning the involute tongues 130 opposite of each other. While the two involute tongues 130 are described as being approximately 180 degrees apart, they may be positioned such that they are not directly opposite. For a typical geometry, the closer the two involute tongues 130 are to directly opposite each other, however, the more the net radial load is reduced.

The geometry of the components of the double involute 132 may impact its performance. For example, changes to the length and shape of the involute channels 134, including the area growth of the channel, may cause changes in the pressure gain in the double involute 132 as well as the loads during operation. Similarly, changes to the final area output and shape of the involute throat 302 may result in changes in the pressure gain of the double involute 132. The double involute 132 may in some embodiments have a configuration designed to optimize pressure gain, minimize net radial loads, or a combination of these or any other factors.

FIG. 4 depicts a flowchart of an embodiment to pump cooling fluid in a cooling system. In one embodiment, one or more components of a centrifugal pump 100 may perform the elements of flowchart 400. In the depicted embodiment, flowchart 400 begins with element 402, receiving cooling fluid into an impeller 106, such as through the fluid entrance 126 of the impeller housing 102. Flowchart 400 may then continue to element 404, where the centrifugal pump 100 drives the impeller 106 to rotate about an axis, such as the pump centerline 110. The centrifugal pump 100 may drive the impeller 106 by, in one embodiment, providing an electric charge to the motor stator 104, which in turn drives the motor magnet 108 of the impeller 106. At element 406, the centrifugal forces generated by the rotating impeller 106 may then force the cooling fluid through fluid channels 114 of the impeller blades 112 to exit the impeller 106.

After the cooling fluid exits the impeller 106, two or more involutes (such as a double involute 132) may receive the cooling fluid at element 408. In one embodiment, the double involute 132 may receive the cooling fluid at two involute tongues 130. The double involute 132 may next, through its geometry, increase the static pressure of the cooling fluid at element 410 by converting the velocity and dynamic pressure of the cooling fluid into static pressure. The double involute 132 may accomplish this by passing the cooling fluid through the involute channels 134 as the area of each involute channel 134 increases. At element 412, any loads, such as a net radial load, generated by the involutes at elements 408 or 410 may be reacted by hydrodynamic bearings such as one or more journal bearings 122 and one or more thrust bearings 124. The involutes may then pass the cooling fluid back into the cooling system (at a higher static pressure), after which the method of flow chart 400 terminates. While the elements of flow chart 400 are shown sequentially, many of the elements may be performed simultaneously. Reacting the generated loads at element 412, for example, may be performed simultaneously and in response to increasing the static pressure of the cooling fluid at element 410.

While certain operations have been described herein relative to a direction such as “above” or “below” it will be understood that the descriptors are relative and that they may be reversed or otherwise changed if the relevant structure(s) were inverted or moved. Therefore, these terms are not intended to be limiting.

It will be apparent to those skilled in the art having the benefit of this disclosure that the present invention contemplates a centrifugal pump with a double involute and hydrodynamic bearings. It is understood that the form of the invention shown and described in the detailed description and the drawings are to be taken merely as examples. It is intended that the following claims be interpreted broadly to embrace all the variations of the example embodiments disclosed.

Although the present invention and some of its advantages have been described in detail for some embodiments, it should be understood that various changes, substitutions and alterations can be made herein without departing from the spirit and scope of the invention as defined by the appended claims. Although an embodiment of the invention may achieve multiple objectives, not every embodiment falling within the scope of the attached claims will achieve every objective. Moreover, the scope of the present application is not intended to be limited to the particular embodiments of the process, machine, manufacture, composition of matter, means, methods and steps described in the specification. As one of ordinary skill in the art will readily appreciate from the disclosure of the present invention, processes, machines, manufacture, compositions of matter, means, methods, or steps, presently existing or later to be developed that perform substantially the same function or achieve substantially the same result as the corresponding embodiments described herein may be utilized according to the present invention. Accordingly, the appended claims are intended to include within their scope such processes, machines, manufacture, compositions of matter, means, methods, or steps. 

1. A centrifugal pump comprising: an impeller housing, the impeller housing comprising a fluid entrance to allow fluid to enter the impeller housing; an impeller located within the impeller housing comprising a plurality of impeller blades, a plurality of fluid channels between the impeller blades, and a motor magnet, wherein the impeller rotates within the impeller housing about a pump centerline in response to an electromagnetic field, and wherein further the fluid channels allow fluid to pass through when the impeller rotates; one or more hydrodynamic bearings positioned between the impeller and the impeller housing to support generated loads; and a double involute coupled with the impeller housing and positioned to receive fluid exiting the plurality of impeller blades.
 2. The centrifugal pump of claim 1, further comprising a motor stator positioned at least partially outside the impeller housing to generate an electromagnetic field in response to an electrical charge.
 3. The centrifugal pump of claim 1, further comprising a motor stator positioned at least partially within the impeller housing to generate an electromagnetic field in response to an electrical charge.
 4. The centrifugal pump of claim 1, wherein the impeller housing is comprised of an injection molded plastic material.
 5. The centrifugal pump of claim 1, wherein the impeller further comprises an impeller body coupled to the motor magnet.
 6. The centrifugal pump of claim 5, wherein the impeller body is positioned between the pump centerline and the motor stator.
 7. The centrifugal pump of claim 5, wherein the motor stator is positioned between the pump centerline and the impeller body.
 8. The centrifugal pump of claim 1, wherein the impeller is at least partially comprised of an injection molded plastic material.
 9. The centrifugal pump of claim 1, wherein the plurality of impeller blades are an even number of impeller blades.
 10. The centrifugal pump of claim 1, wherein the plurality of impeller blades are substantially evenly distributed around the impeller.
 11. The centrifugal pump of claim 1, wherein the one or more hydrodynamic bearings comprise a journal bearing positioned between the impeller housing and the impeller.
 12. The centrifugal pump of claim 1, wherein the one or more hydrodynamic bearings comprise a thrust bearing positioned between the impeller housing and the impeller blades.
 13. The centrifugal pump of claim 1, wherein the double involute further comprises: two involute tongues to receive fluid from the fluid channels; an involute throat to allow exit of fluid from the double involute; and two involute channels to pass fluid from the involute tongues to the involute throat.
 14. The centrifugal pump of claim 13, wherein the two involute tongues are positioned opposite each other.
 15. A method for pumping cooling fluid, the method comprising: receiving cooling fluid into an impeller; driving the impeller to rotate about an axis to force the cooling fluid to exit the impeller; receiving exiting cooling fluid at two or more involutes; increasing a static pressure of the cooling fluid in the two or more involutes; and reacting any loads generated by the two or more involutes and rotating impeller with one or more hydrodynamic bearings.
 16. The method of claim 15, further comprising after increasing the static pressure of the cooling fluid, passing the cooling fluid to a cooling system.
 17. The method of claim 15, wherein driving the impeller to rotate about an axis comprises generating an electromagnetic field to drive the impeller to rotate.
 18. The method of claim 15, wherein increasing the static pressure of the cooling fluid in the two or more involutes comprises passing the cooling fluid through involute channels with increasing area.
 19. The method of claim 15, wherein reacting any loads generated by the two or more involutes and rotating impeller with one or more hydrodynamic bearings comprises reacting any loads generated by the two or more involutes and rotating impeller with a journal bearing.
 20. The method of claim 15, wherein reacting any loads generated by the double involute and rotating impeller with one or more hydrodynamic bearings comprises reacting any loads generated by the two or more involutes and rotating impeller with a thrust bearing.
 21. A centrifugal pump comprising: an impeller housing, the impeller housing comprising a fluid entrance to allow fluid to enter the impeller housing; an impeller means to rotate in response to an electromagnetic field and to allow fluid to pass through during rotation; a hydrodynamic bearing means positioned between the impeller and the impeller housing to support generated loads; and a double involute coupled with the impeller housing and positioned to receive fluid exiting the plurality of impeller blades.
 22. The centrifugal pump of claim 21, further comprising an electromagnetic field generating means to generate an electric field.
 23. The centrifugal pump of claim 21, wherein the double involute further comprises: two involute tongues to receive fluid from the fluid channels; an involute throat to allow exit of fluid from the double involute; and two involute channels to pass fluid from the involute tongues to the involute throat.
 24. A centrifugal pump comprising: an impeller housing, the impeller housing comprising a fluid entrance to allow fluid to enter the impeller housing; an impeller located within the impeller housing comprising a plurality of impeller blades, a plurality of fluid channels between the impeller blades, and a motor magnet, wherein the impeller rotates within the impeller housing about a pump centerline in response to an electromagnetic field, and wherein further the fluid channels allow fluid to pass through when the impeller rotates; one or more journal bearings positioned between the impeller and the impeller housing to support generated loads; one or more thrust bearings positioned between the impeller and the impeller housing to support generated loads; and two or more involutes coupled with the impeller housing and positioned to receive fluid exiting the plurality of impeller blades.
 25. The centrifugal pump of claim 24, further comprising a motor stator positioned at least partially outside the impeller housing to generate an electromagnetic field in response to an electrical charge.
 26. The centrifugal pump of claim 24, further comprising a motor stator positioned at least partially within the impeller housing to generate an electromagnetic field in response to an electrical charge.
 27. The centrifugal pump of claim 24, wherein the impeller is at least partially comprised of an injection molded plastic material.
 28. The centrifugal pump of claim 24, wherein the two or more involutes comprise a double involute, comprising: two involute tongues to receive fluid from the fluid channels; an involute throat to allow exit of fluid from the double involute; and two involute channels to pass fluid from the involute tongues to the involute throat.
 29. The centrifugal pump of claim 28, wherein the two involute tongues are positioned opposite each other. 